There have heretofore been many inventions of heat cycle systems that are capable of converting heat into work or electric power with a high efficiency, using working liquid. For example, JP-A-54-27640 (Japanese Patent Public Disclosure) discloses an electric power generation system that recovers thermal energy of high-temperature exhaust gas. The electric power generation system has a waste heat boiler installed at the upstream side of high-temperature exhaust gas flow path and a fluid preheater at the downstream side thereof. JP-A-61-229905 or GB 2 174 148 A discloses a process of mechanical power generation using a binary cycle comprising a first cycle using two immiscible fluids and a second cycle using a refrigerant fluid. JP-A-2-40007 discloses a power system wherein a reverse Rankine cycle and a Rankine cycle are combined.
First, the thermal efficiency of the heat cycle system will be described below. Assuming that while a working substance is performing one cycle, i.e. undergoing successive changes and then returning to the previous state, it receives a quantity of heat Qh from a high heat source at a temperature Th and loses a quantity of heat Qb from a low heat source at a temperature Tb to do work L (assumed to be a value expressed in terms of heat quantity) to the outside, the following relationship holds:Qh=Qb+L  (Eq. 1)
In heat engines, the work L is given to the outside. In refrigerators or heat pumps, the work L is given to working fluid from the outside. In the case of heat engines, it is desirable that the quantity of heat Qh received from the high heat source should be minimum, and the work L given to the outside should be maximum. Accordingly, the following equation is referred to as thermal efficiency:η=L/Qh  (Eq. 2)
From the above equation, L may be rewritten as follows:η=(Qh−Qb)/Qh  (Eq. 3)
The thermal efficiency η of a heat engine that performs a reversible Carnot cycle may be expressed by using thermodynamic temperatures Th° K. and Tb° K. as follows:η=(Th−Tb)/Th=1−(Tb/Th)  (Eq. 4)
In general, an apparatus that transfers heat from a low-temperature object to a high-temperature object is called a “refrigerator”. The refrigerator is an apparatus that is generally used for the purpose of cooling objects. Meanwhile, an apparatus that transfers heat from a low-temperature object to a high-temperature object to heat the latter is referred to as a “heat pump”. The name “heat pump” may be regarded as an alias for the refrigerator when the usage is changed. The heat pump is used, for example, for a heating operation of an air conditioner for heating and cooling. The relationship between the quantity of heat Qb absorbed from a low-temperature object, the quantity of heat Qh given to a high-temperature object, and the work L (value expressed in terms of heat quantity) done from the outside to operate the heat pump is expressed as follows:Qh=Qb+L  (Eq. 5)
It can be said that, for the same work done, the larger the quantity of heat Qh given, the higher the cost efficiency of the heat pump. Accordingly, the following equation is referred to as the coefficient of performance of the heat pump:ε=Qh/L  (Eq. 6)
From the above Eq. 5, L is:L=Qh−Qb  (Eq. 7)
Hence, the performance coefficient E is expressed as follows:ε=Qh/(Qh−Qb)  (Eq. 8)
Assuming that the absolute temperature of the low heat source is Tb° K. and the absolute temperature of the high heat source is Th° K., a heat pump that performs a reversible Carnot cycle exhibits the largest coefficient of performance among heat pumps operating between the two heat sources. The performance coefficient E of the heat pump is:ε=Tb/(Th−Tb)  (Eq. 9)
The reversible Carnot cycle consists of two isothermal changes and two adiabatic changes and exhibits the maximum thermal efficiency among all cycles operating between the same high and low heat sources.
FIG. 1 is an arrangement plan showing constituent elements of a heat cycle system including a conventional refrigerator J. Refrigerant gas Fg raised in pressure by a compressor C gives heat Qh to a fluid Z in a heat exchanger (condenser) 7, thereby being condensed. Thereafter, the refrigerant is expanded through an expansion valve V. Consequently, the refrigerant lowers in temperature and, at the same time, absorbs heat Qb from a fluid Z2 in a heat exchanger 8 to cool the fluid Z2. Thereafter, the refrigerant is returned to the compressor C and then recirculated. Let us discuss the thermal calculation of a refrigerator arranged as shown in FIG. 1 and adapted to use ammonia as a refrigerant. For the sake of simplicity, let us assume that there is no mechanical loss. The temperature of the refrigerant is 110° C. (T3) at the outlet of the compressor C, 38° C. (T2) at the outlet of the condenser 7, and −10° C. (T) at the outlet of the evaporator V. Therefore, the performance coefficient (theoretically maximum performance coefficient) ε of the refrigerator on the reversible Carnot cycle is:
                                                        ɛ              =                            ⁢                              T                /                                  (                                                            T                      2                                        -                    T                                    )                                                                                                        =                            ⁢                                                                    [                                          273.15                      +                                              (                                                  -                          10                                                )                                                              ]                                    /                                      [                                          38                      -                                              (                                                  -                          10                                                )                                                              ]                                                  ≈                5.4                                                                        (                  Eq          .                                          ⁢          10                )            
In the refrigerator shown in FIG. 1, if the input L (work) of the compressor C is assumed to be 1, the performance coefficient εh of the heat pump (the performance coefficient of the refrigerator +1) is:εh=5.4+1=6.4  (Eq. 11)
FIG. 2 is an arrangement plan showing basic constituent elements of a heat cycle system including a steam turbine (heat engine A). High-temperature and high-pressure steam Fg generated in a boiler B is supplied to a turbine S to rotate it, thereby generating power (work) W. The steam is cooled to form condensate Ee in a condenser Y communicating with the exhaust opening of the turbine. The condensate Ee is raised in pressure by a pump P and then supplied to the boiler B. In the heat cycle system shown in FIG. 2, when waste heat Q2 from the condenser Y is not utilized at all, work W (value expressed in terms of heat quantity) generated from the turbine S is given by:W=Q−Q2  (Eq. 12)
The thermal efficiency ηS of the turbine S is:ηS=(Q−Q2)/Q  (Eq. 13)
In Eq. 12 and Eq. 13, Q is the quantity of heat retained by the working fluid at the turbine inlet side, and Q2 is the quantity of heat output from the working fluid at the turbine outlet side, which is equal to the quantity of waste heat discharged from the condenser Y.
The thermal efficiency η0 of the heat cycle system shown in FIG. 2, i.e. the ratio η0 of work W generated from the turbine S to the quantity of heat (retained heat quantity) Q input to the working fluid in the heat cycle system, is given by:η0=W/Q  (Eq. 14)
If W in Eq. 14 is replaced by W=Q−Q2 of Eq. 12, we have:η0=(Q−Q2)/Q  (Eq. 16)
This is the same as the above-mentioned ηS. Therefore, the following relationship holds:η0=ηS  (Eq. 17)
In the heat cycle system of FIG. 2, if a part or whole Q3 of the waste heat Q2 from the condenser Y is transferred to the condensate at the boiler inlet by a feedwater preheater Y2, i.e.0≦Q3≦Q2  (Eq. 18)and, at the same time, the quantity of heat input to the boiler is reduced by the same amount as the quantity of heat transferred from the condenser Y, then the boiler input heat quantity is Q−Q3. The quantity of heat retained by steam Fg at the inlet of the turbine S is given by:Boiler input heat quantity (Q−Q3)+(heat quantity Q3 transferred by Y2)=Q  (Eq. 19)
The quantity of heat retained by steam Fg at the outlet of the turbine S can be regarded as being Q2. Therefore, power W (value expressed in terms of heat quantity) generated from the turbine S is:W=Q−Q2  (Eq. 20)
Hence, the thermal efficiency ηS of the turbine S is:ηS=(Q−Q2)/Q  (Eq. 21)
Thus, the thermal efficiency ηS of the turbine S is the same as in the case where the waste heat Q2 from the condenser Y is not utilized.
In the heat cycle system of FIG. 2, if a part or whole Q3 of the waste heat Q2 from the condenser Y is transferred to the condensate at the boiler inlet by the feedwater preheater Y2 and, at the same time, the boiler input heat quantity is reduced by the same amount as the heat quantity Q3 transferred from the condenser Y, i.e. (Q−Q3), the thermal efficiency η of the heat cycle system shown in FIG. 2, that is, the ratio of the work W generated from the turbine S, i.e. W=Q−Q2 . . . (Eq. 22), to the input heat quantity of the heat cycle system, i.e. (Q−Q3), is expressed as follows:η=W/(Q−Q3)=(Q−Q2)/(Q−Q3)  (Eq. 23)
In the heat cycle system of FIG. 2, if the waste heat Q2 from the condenser Y is not utilized at all, i.e. Q3=0, the above Eq. 23 becomes as follows:η=(Q−Q2)/Q  (Eq. 24)
In the case of 0≦Q3≦Q2 . . . (Eq. 18), we have:η=(Q−Q2)/(Q−Q3)  (Eq. 25)
In the case of Eq. 25, the denominator is smaller than that in Eq. 24 by −Q3, and hence the value of η becomes correspondingly larger than in Eq. 24.
If the whole Q2 of waste heat from the condenser is transferred to the condensate at the upstream or downstream side of the pump P, we have:Q2=Q3  (Eq. 26)
Hence, the thermal efficiency η of the heat cycle system is:η=1  (Eq. 27)
In the heat cycle system of FIG. 2, the thermal efficiency η thereof in the case of 0≦Q3≦Q2 . . . (Eq. 18) is, as stated above, given by:η=(Q−Q2)/(Q−Q3)  (Eq. 28)
If the denominator and numerator of Eq. 28 are each divided by Q, we have:η=[(Q−Q2)/Q]/[(Q−Q3)/Q]  (Eq. 29)
Eq. 29 may be modified as follows:η=[(Q−Q2)/Q]/[1−(Q3/Q)]  (Eq. 30)
If ηS=(Q−Q2)/Q . . . (Eq. 21) is inserted into Eq. 30, we have:η=ηS/(1−Q3/Q)  (Eq. 32)
In the present invention, even heat having low utility value, such as waste heat, is taken into the heat cycle system by using a heat pump, and a power output is taken out by a turbine in the heat cycle system. The heat cycle system according to the present invention uses heat crossing to extract power from the turbine at high efficiency. When the whole of waste heat Q2 from the condenser Y is utilized, the thermal efficiency η of the heat cycle system is η=1 according to Eq. 27.
As will be understood from the above Eq. 32, the thermal efficiency η of the heat cycle system is determined by the thermal efficiency ηS of the turbine S and the heat quantity Q3 transferred from the waste heat of the condenser Y to the condensate at the upstream or downstream side of the pump P. As Q3 increases to approach Q, the denominator of Eq. 30, i.e. (1−Q3/Q), decreases. Consequently, η increases. It is difficult to increase the heat crossing ratio Q3/Q in heat cycles other than the refrigerating cycle. The reason for this is that it is impossible to increase the temperature difference between a high heat source and a low heat source for heat transfer (heat crossing). Further, Eq. 27 cannot be realized in heat cycles other than the refrigerating cycle.    Patent reference 1: JP-A-54-27640 (Japanese Patent Public Disclosure)    Patent reference 2: JP-A-61-229905 (Japanese Patent Public Disclosure)    Patent reference 3: GB2,174,148A    Patent reference 4: JP-A-2-40007 (Japanese Patent Public Disclosure)